The following terms and their acronyms are used hereinafter in the description of the invention:    Quality—the ratio of the mass of vapor to the total mass of a saturated substance (see G. Van Wylen and R. Sonntag, Fundamentals of Classical Thermodynamics, p.37, (John Wiley & Sons, 2nd ed., SI version1978)).    Flow Quality—the ratio of the vapor mass flow rate to the total mass flow rate.            Access Point (AP)—The device connected to the Internet to transmit data between the Internet and any ED.        End Device (ED)—Any control or monitoring device located remotely from the AP, which sends data or receives commands from the AP.        OCU-ED—An Outdoor Condensing Unit End Device and sensors which monitor performance of a split air-conditioning or heat pump thermal control system.        IAH-ED—An Indoor Air Handler Unit End Device and associated sensors which monitors interior air temperature, and also can monitor performance of the split air conditioner or heat pump thermal control unit.        DHW-ED—A Domestic Hot Water Heater End Device and associated sensors and relays which can monitor and control the hot water temperature being supplied or stored in the hot water tank.        Remote Monitoring System (RMS)—The AP and any ED that communicates with the AP.        Temperature Learning Range—In the currently preferred embodiment, a range of at least 15° F. that will typically occur between 70° F. and 95° F. For example, a data set with data points collected at one degree increments from 72° F. to 87° F. meets the criteria of the Temperature Learning Range.        
The most basic vapor-compression refrigeration system consists of four major components: compressor, evaporator, condenser, and expansion device. Actual practical hardware contains many other critical components for reliable, trouble-free operation, such as a control system, high-pressure and low-pressure safety controls, liquid receiver, accumulator, oil separator, crankcase pressure regulator, etc., but the four basic components are all that is needed to illustrate the function of the bask system and the proposed improvement.
In a typical vapor compression system, refrigerant provides a cooling effect as it evaporates, that is, as it boils and turns from liquid to vapor. For pure refrigerants and azeotropic mixtures, if the refrigerant evaporates at a constant pressure, then evaporation occurs at a constant temperature while both liquid and vapor are present. Likewise, refrigerant rejects energy as it condenses from vapor to liquid. For pure refrigerants and azeotropic mixtures, if the condensation occurs at a constant pressure, then the condensation will occur at a constant temperature until all the vapor has condensed to a liquid. Therefore, for evaporation or condensation, the temperature and pressure are related by what is known as the pressure/temperature saturation curve.
In the conventional basic vapor compression cycle shown schematically in FIG. 1, subcooled liquid refrigerant 111 leaves the condenser 112 at high pressure and flows to the throttling device 113 (capillary tube, TXV, etc.) where the pressure is decreased. The refrigerant then enters the evaporator 114 as a two-phase mixture (liquid and vapor) and evaporates or boils at low temperature, thereby absorbing heat. Superheated refrigerant vapor 115 exits the evaporator and enters an electrically-driven compressor 116 where the pressure and temperature are increased as the compressor compresses the refrigerant vapor. The refrigerant vapor leaving the compressor 117 is superheated, and this refrigerant is cooled and condensed in the condenser 112 where heat is rejected, and the refrigerant is condensed to liquid. Refrigerant typically leaves the evaporator 114 slightly superheated (superheat vapor) to assure evaporation has been complete. Refrigerant typically leaves the condenser 112 slightly subcooled (subcooled liquid) to assure condensation has been completed.
FIG. 2 schematically shows vapor compression system that is known as a “split” air conditioning system which is well known in the art. It includes an outdoor unit 20, also referred to as a condensing unit, and an indoor unit 40, also referred to as the fan coil unit or indoor air handler, housed in a structure such as a residential home or a commercial building. The outside of the structure is denoted by numeral 37 and the inside of the structure by numeral 38 with the structure's exterior wall being denoted by numeral 39. The indoor (40) and outdoor (20) units are plumbed together via a liquid line 31 and a vapor line 35 in a known way. A third approximately atmospheric pressure condensed water drain line 36, also known as a condensate drain line, carries water 44 that condenses on the evaporator 42 and is captured in the drain pain 45 of the indoor unit 40. This condensate water is then transported to the outside 37 of the structure, typically to be deposited on the ground 26 near the outdoor unit 20 to create moist wet soil 54. The condensate drain line 36 carries the condensed water from inside 38 being cooled by the system to the outside 37 of the structure (typically by gravity assisted flow only), and is typically bundled with the two refrigerant pipes 31,35 along with any control wires connecting the controls of the indoor and outdoor unit (not shown).
The split system outdoor unit 20 typically includes a compressor 22, condensing coil 23, and cooling fan/motor unit 24 as well as other components well known in the art and is typically located on a concrete or plastic slab or foundation 25 that rests on the ground 26. Standard operation during the cooling season consists of superheated refrigerant vapor entering the condensing unit 20 via the vapor refrigerant line 35. The flow path consists of passing through the compressor 22 and condenser 23, and exiting the outdoor unit 20 via liquid refrigerant conduit 31 and flowing to the indoor unit 40. The cooling fan/motor 24 provides air flow across the condenser. The indoor unit consists of a throttling device 41 (such as a thermal expansion device, orifice plate or capillary tube), evaporator 42 and blower 43 as well as other components well known in the art. Subcooled liquid refrigerant enters the indoor unit 40, flowing to the expansion throttling device 41 and then the evaporator 42 and normally exiting the evaporator as superheated refrigerant vapor flowing back to the inlet of the compressor 22 (which is located in the outdoor unit 20) through vapor refrigerant line 35. Condensate, i.e. condensed water 44, flows by gravity (or by being pumped) in the condensate drain line 36 and exits onto the ground 26 in a region near the outdoor unit 10. Those skilled in the art will understand the foregoing is only a very brief discussion of the operation of a split vapor compression air conditioning system for purpose of defining the condensing unit 20 and the purpose of the liquid line 31, compressor 22, vapor line 35, condenser 23, expansion device 41 and the evaporator 44 in such a system.
It is also well known in the art to use a reversing valve in the outdoor unit 20 along with check valves and two expansion valves to configure the vapor compression system into a device that provides both cooling during warm ambient temperatures and heating during cold ambient periods. Such a vapor compression system, with a reversing valve, is commonly referred to as a heat pump. For both a split air conditioning unit and a split heat pump there are two refrigerant lines connecting the units, one containing condensed liquid refrigerant the other containing vapor refrigerant. In both cases, refrigerant vapor exiting the evaporator flows to the compressor inlet and condensed refrigerant leaving the condenser flows to the throttling valve then on to the evaporator as more simply depicted in FIG. 1.
In a typical residential home, one of the largest sources of energy consumption is the split vapor-compression air conditioning or heat pump system described above (also hereafter referred to as the A/C unit). If the A/C unit is operating at degraded efficiency, an equipment owner may be unaware of the inefficiency because the equipment operates at a higher duty cycle to maintain the house at the appropriate temperature. Eventually, an undetected equipment problem might lead to a costly system failure, such as a compressor motor burnout, or simply lead to increased energy consumption, and failure to cool living spaces adequately during the next hot day.
A/C units that fail on the first unseasonably hot day of the year have typically been operating at reduced capacity (at high duty cycles) for weeks or months. Although the energy bill is higher due to this degraded performance, equipment owners either fail to notice the energy increase or fail to relate the high energy bill to the A/C unit's degraded efficiency. An RMS that can detect degraded A/C unit performance prior to equipment damage or long periods of inefficient operation has obvious benefits to the equipment owner, electric utility, and the environment if such a unit is economical and reliable.
The HVAC service provider also benefits from an RMS by distributing service calls more evenly. Currently, when the first hot days arrive and the unit's degraded capacity becomes apparent, the equipment owner will call a HVAC technician to service the A/C unit. However, the service provider is usually overwhelmed with similar calls on the first hot day where system degradation becomes obvious. For the largo service contractor firms with tens of thousands of service contracts, an RMS that can predict failures before they occur or determine inefficient operation has important benefits. The RMS also assists in the scheduling of service calls by providing detailed information regarding the severity of the problem and by offering remote diagnosis. This allows the service provider to dispatch the appropriate repair technician to the site and cluster similar service calls to a technician specialized in that repair.
There are many complex monitoring system approaches, such as the one disclosed in U.S. Pat. No. 7,469,546 that includes the use of temperature, pressure and flow sensors. Using pressure and flow sensors dramatically increases the cost of the monitoring system and makes this type of monitoring system unfeasible for residential or small commercial A/C units. A cost-effective RMS for residential or small commercial A/C units must eliminate expensive sensors and the labor-intensive process of plumbing the sensors into the refrigeration flow circuit. Flow meters and pressure transducers with sufficient accuracies that are included with the complex monitoring system are too expensive for this application. To reduce costs to practical levels, the temperature sensors that directly measure refrigerant temperatures (by being plumbed directly into the refrigerant flow loop) must also be replaced by exterior, surface mount temperature sensors. The in-loop temperature sensors are significantly more expensive than surface mount sensors due to the increased cost of a refrigerant-compatible and pressure-tolerant sensor and the additional cost to install these temperature sensors into the refrigeration flow circuit.
In addition, for the typical split A/C unit or heat pump system, where the evaporator and blower are located inside the conditioned building (and referred to as the air handler or evaporator section) and the condenser and compressor are located outside the conditioned space (and referred to as the condensing unit), it is more difficult to monitor refrigerant pressures, temperatures or flow rates on both the evaporator section and the condenser unit locations since they are not collocated.
U.S. Pat. No. 6,385,510 discusses a remote monitoring method where the conditioned air return temperature and return air humidity, along with the supply air temperature and the system's design airflow rate and rated cooling capacity are used to monitor performance. Using the system's designed airflow rate can, however, introduce a significant amount of error as the air handler airflow rate is a function of the pressure differential across the blower. The geometry, length, or circuiting of the air supply and return ducting will not be identical at all installations, altering the pressure differential across the blower and, therefore, the blower airflow rate. An even larger inaccuracy could be created by the air filter, where there is a wide assortment of air filters that impose different pressure drops and where pressure drop across a filter will change over time due to the filter collecting particles. If the air filter has become significantly clogged, the air handler airflow rate decreases and then this monitoring system would think that cooling capacity was enhanced due to the higher air enthalpy change when, in reality, cooling capacity has been diminished by a clogged air filter.
It is also well known in the art to have a server that is capable of communicating with one or more remote locations, to send and receive data from these remote units for the purpose of monitoring or controlling multiple devices. For example, U.S. Pat. No. 7,792,256 discloses a system for remotely monitoring, controlling, and managing one or more remote premises.
We have discovered that there is a far more cost-effective way to do remote monitoring using compressor discharge sampling without sacrificing accuracy in predicting the health of the system being monitored. In order to appreciate just how significant our discovery is, however, some additional background discussion is useful.
It well known to those skilled in the art that when a fluid, such as refrigerant in the evaporator of a vapor compression system of the types shown in FIGS. 1 and 2, evaporates, that refrigerant does so at a constant temperature as long as the pressure is constant. From a mathematical perspective, the evaporating pressure and evaporating temperature are related and not independent. That is, for a specific refrigerant, if the evaporating pressure is known, then the evaporating temperature can be determined by reference to that refrigerant's known saturation pressure temperature relationship which is valid only when the refrigerant is saturated. Once all the refrigerant liquid has evaporated, additional heat input will cause the refrigerant temperature to increase above saturation temperature and the refrigerant is referred to as being superheated. The numerical increase in the temperature of the refrigerant above the saturation temperature is referred to as the superheat of the refrigerant. That is the mathematical difference between the actual temperature and the saturation temperature (at that pressure) is the superheat of the refrigerant.
For example, using the NIST Standard Reference Database 23, version 8.0 software program titled REFPROP, available from the US Department of Commerce; the saturation temperature of Refrigerant 134a at 40 psia is 29 degrees F. If the refrigerant has been heated to 34 degrees F., (and the pressure held at 40 psia) then the superheat would be 5 degrees F.
Therefore, we recognized when examining the condition of the refrigerant exiting the evaporator, with only a temperature sensor, the diagnostic value of the knowing both the temperature and pressure at this exit can be undesirably limited. If the temperature is above the saturation temperature, as in the case just discussed when the refrigerant is superheated, then the exit enthalpy, or similarly the thermodynamic state point, can be determined, However, if the temperature exiting the evaporator is the saturation temperature (for the measured pressure), then it is not possible to determine the flow quality from only these two measurements. That it, is not possible to determine if the refrigerant flowing from the evaporator outlet is nearly all liquid, nearly all vapor, or some other saturated condition. Without this knowledge of the flow quality, the thermodynamic state point or the exit enthalpy of the saturated refrigerant exiting the evaporator outlet cannot be determined. This also means the effectiveness of the evaporator at evaporating the entire refrigerant, cannot be determined with only the exit pressure and temperature measurements, if refrigerant is leaving the evaporator in a saturated condition. It is for this reason, that thermal expansion valves and other feedback expansion devices are necessary to monitor exit superheat, and the evaporators of these systems are designed to have superheated refrigerant exit the evaporator. This exit superheat is typically very small, since the primary function of the evaporator is to vaporize refrigerant. Under certain fault conditions, the superheat may be excessive, such as when the refrigerant charge is low, in other cases the superheat may be negligible or the evaporator exit may be saturated. When the refrigerant exits the evaporator as a saturated refrigerant, it is not possible to determine the quality that is the relative amount of vapor and liquid refrigerant exiting the evaporator; hence, it is also not possible to determine how much the cooling capacity of the evaporator has been degraded.
The foregoing point is most easily demonstrated by the following example. For a unit mass of refrigerant, the heat absorbed by evaporation, i.e., the heat of vaporization is much greater that the heat capacity of a single phase superheated refrigerant. In the case, for example, of above-mentioned Refrigerant 134a and using REFPROP once again, the energy to evaporate one pound of Refrigerant 134a from saturated liquid to saturated vapor (at a saturation pressure of 40 psia, saturation temperature of 29 degrees F.), namely the latent heat of vaporization is 86.0 BTU/LB, whereas the energy to raise the temperature of the refrigerant from the saturation temperature of 29 degrees F. to 34 degrees (5 degrees F. of superheat) is only 1.1 BTU/LB or only 1.3% of the latent heat of vaporization (1.1/86). Likewise, the energy to raise the temperature of the refrigerant from the saturation temperature of 29 degrees F. to 39 degrees (10 degrees F. of superheat) is only 2.1 BTU/LB or 2.4% (2.1/86). Therefore if the cooling provided by evaporation increased by about 1% (energy into the evaporator increased by 1%) the superheat would increase from 5 degrees F. to about 10 degrees F., and therefore this increase could be easily determined by a noticeable temperature change of 5 degrees F.
If, however, the cooling provided by evaporation decreased by anything more than about 1% (energy into the evaporator decreased by more than 1%) there would be no superheat at the exit; rather the refrigerant would be leaving the evaporator at saturated conditions, in the case of this example, 29 degrees F. The two important points to be made here are (1) by simply monitoring the temperature in the conventional way it is not possible to determine if the cooling (namely the evaporation in the evaporator) has decreased by 1% or 90%, since the refrigerant exiting the system would be saturated and therefore at the same temperature, and (2) a very accurate measurement of the temperature at the outlet of the evaporator is necessary if one has any hope of determining any reduction in cooling capacity, since a 5 degree reduction in outlet temperature is only a 1% reduction in capacity and, once saturated outlet conditions are achieved, no further temperature measurements are useful.
We have discovered an inexpensive diagnostic method that makes it possible to identify reductions in cooling before they become significant and to provide this diagnostic warning at minimal cost. To reduce cost, all measurements are collected at a single location, namely the condensing unit, without the need to install sensors inside the structure being cooled or on the indoor air handling unit. While our method could measure the refrigerant temperature at the inlet to the compressor, which represents the first component downstream of the outlet of the evaporator, or measure the temperature anywhere in the refrigerant line between the evaporator outlet and the compressor inlet, we recognized that even small heat transfer with the ambient could affect an accurate reading of the superheat temperature and if incomplete evaporation were occurring, the same inability to determine the flow quality, with only a temperature measurement would still be present.
We have discovered another and far superior method of determining a reduction in the evaporation at the outlet of the evaporator that only relies on a temperature measurement, but can still determine the relative amount of saturated vapor, that is the relative quality, exiting the evaporator. Since, as we recognized, the compressor normally inputs a relative constant amount of energy into the refrigerant (for a specific outdoor air temperature), and the refrigerant always exits the compressor as a superheated vapor, by investigating the temperature at the outlet of the compressor, that is by looking at the compressor discharge temperature, the relative amount of evaporation in the evaporator can be easily determined. If the refrigerant exits the evaporator with a low thermodynamic flow quality, meaning a large fraction of saturated liquid is leaving the evaporator, the temperature at the outlet of the compressor will be lower. Likewise if the refrigerant exits the evaporator with a high thermodynamic flow quality, meaning a small fraction of saturated liquid is leaving the evaporator, the temperature rise at the outlet of the compressor will be far greater, and if the refrigerant exits the evaporator as a superheated vapor, the temperature rise at the outlet of the compressor will be even more. Once again an example from REFPROP may be useful.
If 1 pound per hour of R-134a refrigerant (at 40 psia) leaves the evaporator with a quality of 0.5 (half vapor by mass) which represents an exit enthalpy of 128.1 BTU/lb and has the compressor input 60 BTU/hr of energy into the system, the refrigerant will be discharged from the compressor with an enthalpy of 188.1 BTU/Lb. If the pressure at the compressor discharge is 140 psia, then the refrigerant is discharging the compressor at a temperature of 131.6 degrees F.
Alternatively, if 1 pound per hour of R-134a refrigerant (at 40 psia) leaves the evaporator with a quality of 0.9 (90% vapor by mass), which represents an exit enthalpy of 162.5 BTU/lb and has the same compressor energy input of 60 BTU/hr into the system, the refrigerant will discharge the compressor with an enthalpy of 222.5 BTU/Lb. If the pressure at the compressor discharge is again 140 psia, then the refrigerant is now being discharged from the compressor at a temperature of 269.5 degrees F. A 137.9 degree F. increase occurs due to the quality change from 0.5 to 0.9.
Finally, if 1 pound per hour of R-134a refrigerant (at 40 psia) leaves the evaporator with a superheat of 5 degrees F. (exit temperature of 34 degrees F. and pressure of 40 psia), which represents an exit enthalpy of 172.1 BTU/lb and has the compressor inputs the same 60 BTU/hr of energy into the system, the refrigerant will be discharged from the compressor with an enthalpy of 232.1 BTU/Lb. If the pressure at the compressor discharge is again 140 psia, then the refrigerant is now being discharged from the compressor at a temperature of 306 degrees F. A 36.5 degree F. increase due to a change from a saturated at a quality of 0.9 to superheated 5 degrees F.
As the foregoing example clearly demonstrates, by measuring the temperature at the compressor discharge, rather than the evaporator inlet, the effects of different evaporator exit conditions, is greatly amplified. This simplifies the measurement and less accurate temperature measurements on the exterior of the tube can be used since the temperature differences are much larger than the typical 5 degree F. variation at the evaporator outlet (between superheated evaporator discharge and a saturated discharge). Also by measuring the temperature at the compressor discharge, different evaporator exit qualities, i.e., different levels of evaporation can also be determined from the corresponding compressor discharge temperature.
Hence, one object of the present invention is to provide a reliable, low-cost vapor-compression air-conditioning and heat pump monitoring system that can reliably predict equipment failures and determine low-efficiency operation. This monitoring system can be located entirely in the outdoor condensing unit, without the need to place any sensors inside the building or inside the indoor portion of the split air conditioning or heat pump unit. This greatly simplifies installation and lowers cost.
The present invention disclosed herein utilizes only three temperature sensors, a current sensor, and three voltage sensors. Two of the temperature sensors are mounted on the exterior surfaces of pipes located in the outdoor condensing unit, while the last temperature sensor measures outdoor air temperature. The tube-mounted temperature sensors are located on the compressor discharge tube and the condenser outlet tube. The current draw of the compressor or the total current to the condensing unit, which includes the power to both the compressor and the condenser fan, is also measured to identify potential problems. Two voltage measurements (Run Winding Voltage, Start Winding Voltage), are also used. Optionally, the inlet voltage can also be measured, to verify the status of the condensing unit's contactor (relay). With this information, the system, according to the present invention, can automatically learn the characteristics of the specific vapor compression system, then monitor the system for future problems, including faulty run or start capacitance operation, low refrigerant charge, reduced condenser airflow or reduced evaporator airflow. In addition to split air conditioning or heat pump systems, this same diagnostic approach can be applied to any vapor-compression system, including refrigerators, freezers and the like.
We have developed a family of novel diagnostic algorithms to enable the identification of all common mechanical problems, electrical problems, and maintenance issues. These algorithms have been designed to be very simple, thereby allowing the analysis to be performed on site or at a remote location, by transferring the data via the internet or other means, or by using a combination of both on-site and remote analysis to allow reduced data traffic, safe storage of the data, and reduced server loading.
In one currently preferred embodiment, the system consists of an indoor AP device which provides communication between the OCU-ED and the server computer, located at a remote location and connected via the Internet or other appropriate communications system. The OCU-ED can be installed quickly on the condensing unit without any modifications to the plumbing or condensing unit since measurements are obtained on external tube surfaces in the outdoor condensing unit, the ambient outdoor air, from voltage measurements from capacitors and contactor connection surfaces in the outdoor unit, or from current draw of a compressor or overall outdoor condensing unit power supply conductor either of which is routed through a current sensor.
The AP device of the present invention would also allow additional end devices to be located in the building. For example, a second end device, referred to here as the IAH-ED, could be used to monitor the indoor air temperature returning to the air handler and can therefore provide a thermostatic like thermal control effect, to override the existing thermostat and provide precise temperature control based on instructions from the AP. That is, a programmable thermostatic effect, selected via the Web-based interface and transmitted by the AP to this end device, could be used to control the room air temperature. Likewise, one skilled in the art could extend this invention to other items in the home, such as for example a DHW-ED, which along with associated sensor and relay, can monitor and control the domestic hot water supply temperature based on instructions from the AP. That is a programmable time-dependent thermostatically-controlled hot water heater effect, selected via the Web-based interface and transmitted by the AP to this DHW-ED, could be used to monitor and control the domestic hot water supply temperature.